Hydrostatic transmission

ABSTRACT

An improved compact design for a hydrostatic transmission having a hydraulic pump and hydraulic motor mounted on a center section in a housing, wherein the pump and motor are mounted at generally right angles to one another, and the longitudinal axis of the input shaft is located between a first and second parallel planes located at respective ends of the motor shaft and the longitudinal axis of the motor shaft is located between a third and fourth parallel planes located at respective ends of the pump shaft.

This application is a continuation of U.S. patent application Ser. No.10/702,333 filed Nov. 6, 2003 now U.S. Pat. No. 6,941,753, which is acontinuation of U.S. patent application Ser. No. 10/243,368 filed Sep.13, 2002 now U.S. Pat. No. 6,672,057, which is continuation of U.S.patent application Ser. No. 09/880,587 (U.S. Pat. No. 6,550,243) filedJun. 13, 2001, which is a continuation of U.S. patent application Ser.No. 09/846,545, filed on May 1, 2001 now abandoned, which is acontinuation of Ser. No. 09/420,183 (U.S. Pat. No. 6,256,988), filedOct. 18, 1999, which is a continuation of Ser. No. 09/016,584 (U.S. Pat.No. 6,014,861), filed Jan. 30, 1998, which is a continuation of Ser. No.08/644,474 (U.S. Pat. No. 5,768,892), filed May 10, 1996, which is acontinuation of Ser. No. 08/613,371 (U.S. Pat. No. 5,616,092), filedMar. 11, 1996, which is a continuation of Ser. No. 08/260,807 (U.S. Pat.No. 5,501,640), filed Jun. 16, 1994, which is a continuation of Ser. No.08/025,272 (U.S. Pat. No. 5,330,394), filed Mar. 2, 1993, which is adivision of Ser. No. 07/917,858 (U.S. Pat. No. 5,314,387), filed Jul.22, 1992, which is a continuation-in-part of Ser. No. 07/727,463 (U.S.Pat. No. 5,201,692), filed Jul. 9, 1991. All of these prior applicationsare incorporated herein by reference.

BACKGROUND OF THE INVENTION

This invention relates generally to transaxles including a hydrostatictransmission (“HST”) commonly used with riding lawn mowers and similarsmall tractors. Such tractors generally use an engine having a verticaloutput shaft which is connected to the transaxle via a conventional beltand pulley system. A standard HST for such a transaxle includes ahydraulic pump, which is driven by the engine output shaft, and ahydraulic motor, both of which are usually mounted on a center section.Rotation of the pump by an input shaft creates an axial motion of thepump pistons. The oil pressure created by this axial motion ischannelled via porting to the hydraulic motor, where it is received bythe motor pistons, and the axial motion of these pistons against athrust bearing causes the motor to rotate. The hydraulic motor in turnhas an output shaft which drives the vehicle axles through differentialgearing.

Among the advantages of transaxles with hydrostatic transmissions arethe reduction of the number of parts and in the size of the unit, and,in some instances, the elimination of mechanical gears. As is known inthe art, the use of a transaxle having a hydrostatic transmissionenables the manufacturer to include all necessary elements in one unit,whereby the transaxle is easily incorporated into the tractor design, asit requires only the addition of a belt to connect it to the motor and acontrol lever for changing speed and direction. While the basicprinciples of transaxles with an HST are well known in the prior art,there are several disadvantages of present transaxles with HST designs.These disadvantages, and the present invention's means for overcomingthem, are set forth herein.

A major problem with some prior transaxle designs is that thetransmission is too large and too expensive to be used with the smallertractors where it would be most effective. An attempt to solve thisproblem is shown in Okada, U.S. Pat. Nos. 4,914,907 and 4,932,209. TheOkada '209 patent discloses a first mechanical deceleration means,namely the gear on the motor shaft and countershaft within the axlehousing, and a second mechanical deceleration means in the differential.The gearing in the deceleration means eventually transmits power to thedifferential gears, which are then used to drive the output axle.However, these mechanical deceleration units add unnecessary weight andexpense to the unit. An object of the present invention is to provide antransaxle design which does not require such additional mechanicaldeceleration means.

Another variation on the standard transaxle with HST design is shown inThoma, U.S. Pat. No. 4,979,583. This patent teaches the segregation ofthe hydraulic units from the remaining portions of the transaxle throughthe use of separate segregated cavities to house each. In addition, thepump and motor in the Thoma design are mounted back-to-back, so that theinput and output shafts have the same orientation. Thus additional gearunits are required to re-orient the rotation of the output shaft so thatit is parallel to the ultimate drive axle. Further gears then drive adifferential which rotates the drive axle. This additional gearing addsweight to the unit and expense to the manufacturing process.

Thus, the Okada and Thoma designs present problems from the standpointof manufacturing a small, economical transaxle including an HST which iseasily adaptable to different size tractors or axle configuration. Okadarequires multiple gearing and Thoma requires a housing having segregatedcavities. The present invention is designed to overcome these and otherproblems in the prior art by providing a compact, economical transaxlewith HST which substantially reduces the number of moving partspreviously required.

SUMMARY OF THE INVENTION

The present invention, sometimes referred to generally as a “transaxle,”includes a split-axle housing which encases an HST. The HST includes apump and a motor whose orientation to one another may be variedaccording to the space requirements dictated by the size andconfiguration of the vehicle. This transaxle also includes a novelhydraulic reduction means, an improved differential, a longer lasting,more effective means of preventing oil leakage from the axle shafts inthe housing, a center section supporting the output drive shaft, animproved means for hydraulically bypassing the HST and a unique checkvalve arrangement. Each of the specific novel improvements are combinedto provide a transaxle which is compact, reliable and economical tomanufacture. These and other objects and improvements of this inventionwill be set forth in more detail herein.

One object of this invention is to provide an improved transaxle whereinthe center section of the HST, on which the pump and motor are mounted,also serves as the bearing support of the output drive shaft. In theprior art, for example, Okada U.S. Pat. No. 4,932,209, one end of thegear drive arrangement is supported in the center section, but the otherend is supported by the upper and lower axle housing casings.

The advantage of the present invention's arrangement is that iteliminates the need for an additional bearing support, thus reducing thecosts and assembly time required. It also eliminates the toleranceconcerns for aligning the bearing supports for the output drive shaft.

A further object of this invention is to provide a transaxle that mayuse multiple mechanical reduction units, but requires only a single suchunit because a portion of the overall reduction is providedhydrostatically. The prior art generally requires dual or multiplemechanical reduction units in conjunction with the hydraulic unit. Forexample, as set forth above, U.S. Pat. No. 4,932,209 requires the use oftwo separate mechanical reduction units, including a separatecounter-shaft between the hydraulic motor and the differential used todrive the output axle.

The present invention makes this same reduction through the hydraulicsitself by the use of a motor which is larger in displacement than thepump. This eliminates the need for any secondary mechanical reductionunits, thereby reducing sources of possible mechanical failure. Thesingle reduction arrangement reduces the number of necessary componentsand the size of the transmission, and it eliminates the need for anadditional support shaft or jack shafts, thus resulting in a smaller,simpler and less expensive transaxle. In a heavy duty application, theprior art often used two sets of mechanical reduction units to handlethe necessary reduction. In such instances, the present invention'shydraulic reduction can eliminate the need for such multiple reductionunits or could be used in conjunction with secondary units only.

A further object of this invention is to restrict the oil from having toextend to the outer axle support bearings, as is common in prior artmodels. The gearing and the hydrostatic transmission element of thisinvention are enclosed in a single chamber formed by an upper casing anda lower casing. The axle shafts extend through this chamber and aresupported by separate bearing surfaces outside of the chamber.

In most of the prior art, the entire axle casing is filled with oil outto the outer axle bearings to provide lubrication to these bearings, inaddition to the hydrostatic function of the oil in the pump and motor.However, after the outer axle bearings wear through use, theeccentricity or “play” in the shaft may distort the oil seal at saidouter bearings, allowing the leakage of oil out of the main chamber.Maintenance of a leak-free joint is critical to the function andappearance of such a transaxle with HST unit. The entire internalhydraulic parts of an HST should be covered with oil, as an insufficientamount of oil in the main transmission cavity will cause foaming of theoil, damaging the hydraulic structures. Excessive oil leakage is aserious problem as it will hamper the ability of the HST to operate andcause damage to the internal workings of the HST. Oil leakage alsopresents an aesthetic problem for manufacturers of transaxles, ascustomers are usually quite disturbed by the presence of oil leaks andthe accompanying oil stains. Thus, the reduction or elimination of oilleakage is critical for the continued success of transaxle sales.

In the prior art, maintenance of such a leak-free joint at the outerbearings requires the use of extra bolts and sealant, which addadditional weight and cost to the unit. An additional problem with priorart designs is that such wear in the outer axle bearings can also causecontamination of the oil due to the presence of “shavings” and otherdetritus from the worn bearings.

Although such construction could be used with the other novel elementsof the present invention, to solve these problems of leakage andpotential oil contamination at minimum cost, the present invention alsopresents a unique means of restricting the oil to those portions of thetransaxle where it is needed to lubricate the differential and to workthe pump and motor of the HST. Thus, chambers separate from the mainchamber enclosing the HST and differential surround the majority of eachaxle shaft. Therefore, the oil does not extend throughout the entirecasing or to the outer axle bearings, removing the potential problem ofoil leaking from the casing. Separate grease pockets are used tolubricate these outer axle bearings, resulting in a much more durableseal and allowing for the use of a higher viscosity grease lubricatethese outer axle bearings.

This improvement also allows for a reduction in the amount of oil neededto fill the transmission case, and, due to the reduced sealant area atthe outer axle bearings, a reduction in the amount of sealant required.Due to the fact that the maintenance of a leak-free joint at the outeraxle bearings is not required, this invention also allows for reducedmanufacturing tolerances, which reduces the manufacturing costs of theunit.

A further improvement is in the method used to place the transaxle intoneutral gear to enable movement of the tractor without the motorrunning. A problem with the typical HST arrangement is that “neutralgear” does not exist, as it is merely a point where the hydraulicpressure in the pump goes to zero. However, at this point the oilremains in the transmission, preventing the vehicle from being rolledfreely.

The prior art generally solves this problem by diverting the oil througha hydraulic valve from the pressure side to the vacuum side of the HSTcenter section. The problem with such a design is that the hydraulicvalve allows for the movement of only a limited amount of oil due toinherent design limitations, such as the diameter of the hydraulic valuethrough which the oil is diverted. Furthermore, machining such a valverequires precise tolerances, thus increasing the manufacturing costs ofthe unit.

In the present invention, this problem is solved by providing amechanism whereby the motor block is mechanically lifted from itsrunning surface, thereby allowing the oil to bypass the vacuum-pressurecircuit and to exit the case completely. This operates to enable thevehicle to freewheel more easily than is possible with the prior arthydraulic valve method.

Another object of the present invention is to provide an improved designof the motor and motor thrust bearing in a hydrostatic transmission,whereby the motor shaft does not extend through the motor thrustbearing, and thus the bearing is fully supported and does not require anintermediate support plate, as is used on prior art models.

For example, U.S. Pat. No. 4,953,426 to Johnson teaches a thrust bearinghaving a motor shaft extending through its center section. As in thepresent invention, the thrust bearing in patent '426 is supported by onesection of the housing. However, because the '426 thrust bearing has themotor shaft extending through its center, it is not solely supported bythe housing, but rather is supported by two “fingers” on either side ofthe thrust bearing. To support the thrust bearing against the hydraulicforces applied by the motor pistons, an additional structurallysignificant piece is required to support between these fingers.

In the present invention, the thrust bearing is fully supported by thehousing part into which it is inserted, thus eliminating the need for anadditional structural member. This results in an assembly that issimpler and less expensive to manufacture.

A further object of this invention is an improved differential gearassembly. In the prior art, differential assemblies generally require across-shaft to support the planet bevel gears. The arrangement of thepresent invention eliminates the need to use such a cross-shaft byproviding a simple end cap axle support and bevel and planet gearentrapment.

A further novel feature of this invention is in the placement of thebrake portion in the housing. Disk brakes are known in the art, andgenerally consist of a series of disks or plates, mounted on or about arotating shaft, with at least some of the disks or plates rotating withthe shaft. Such brakes generally have a brake arm or level which ismoved to activate the braking feature by a means for transmitting themovement of the brake arm to the series of disks, causing the stationarydisks to be pressed against the rotating disks, thus braking thisrotating shaft through friction. This means for transmitting themovement of the brake arm to the disks generally consists of rods orshafts, and, in the prior art, these rods or shafts were mounted in ahousing which is separate from the housing containing the HST. In thepresent invention, the brake rods are mounted directly into the HSThousing through half-round sections formed into each of the matinghousing sections, thus eliminating the need for this separate housingand reducing the manufacturing costs of the products.

An additional novel feature of this invention is the design of the checkvalve for the center section. Prior art check valve designs generallyuse hardened steel balls working against a steel or cast iron seat. Tominimize the overall weight of the transaxle unit, however, the centersection of the present invention is preferably made of cast aluminum,which is not strong enough to function as such a valve seat and towithstand the wear from such a check valve operation. This problem issolved by the use of a steel insert in the center section to support thesteel balls.

To create a seal at such a location, it is known to use a machinedsurface on both the seat and the insert, so that a standard O-ring sealcould be used. However, use of such a sealing means would requireadditional machining steps on the seat and insert, adding to the overallmanufacturing costs of the unit.

To overcome these problems in the prior art, the present invention callsfor the use of a powdered metal plate which acts as both the check valveseat and as the seal. The sealing functions of the plate are createdthrough the use of a raised surface on the plate, which is pressed intothe lower strength aluminum to form a seal. This design has theadvantage of being simple and inexpensive to manufacture, whilemaintaining the advantage of a light overall weight.

It is a further object of this invention to provide an improvedhydrostatic transmission wherein the pump and the motor of the HST neednot be orientated at a 90-degree angle to one another as required by theprior art. In the present invention, the 90-degree orientation is thepreferred embodiment. However, an orientation other than 90-degrees canbe achieved by use of a helical gear between the output drive shaft andthe differential.

Further explanation and details of the above objects of this invention,as well as other benefits and advantages of this invention, will be setforth in the following sections.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a top view of a transaxle with a hydrostatic transmissionmanufactured in accordance with this invention.

FIG. 2 is a fragmentary elevational section view along the line B—B ofFIG. 1.

FIG. 3 is a fragmentary side view along the line C—C of FIG. 1.

FIG. 4 is a fragmentary elevational section view along the line A—A ofFIG. 1.

FIG. 5 is a perspective view of a center section of a hydrostatictransmission in accordance with this invention.

FIG. 6 is a bottom plan view of a center section of a hydrostatictransmission in accordance with this invention.

FIG. 7 is the top view of the check valve plate for the center sectionof a hydrostatic transmission.

FIG. 8 is a fragmentary section view along the line D—D of FIG. 7.

FIG. 9 is a detailed view of the motor and motor shaft of thehydrostatic transmission of the present invention.

FIG. 10 is a top view of the endcap of the differential of the presentinvention.

FIG. 11 is a side view of the endcap of FIG. 10.

FIG. 12 is an end view of the endcap of FIG. 10.

FIG. 13 is a top view of a ring gear used in the differential of thepresent invention.

FIG. 14 is a top view of one embodiment of the entire differential.

FIG. 15 is a top view of an end cap used in one embodiment of thedifferential.

FIG. 16 is a sectional view of the end cap used in the differentialshown in FIG. 14 the E—E axis in FIG. 15.

FIG. 17 is a side view of the planet gear used in the differential.

FIG. 18 is a side view of the bevel gear used in the differential.

FIG. 19 is a top view of a portion of a transaxle showing the brakingmechanism used with this invention.

DETAILED DESCRIPTION OF THE DRAWINGS

All hydrostatic transmissions operate on the principle of an input shaftdriving a pump, which, through the action of its pistons, pushes oil toa motor, which rotates a motor shaft. This rotation is eventuallytransferred through a differential gearing system to drive an axleshaft. With these general principles in mind, we turn to the drawings ofthe present invention showing the various improvements made by thisinvention on the prior art.

FIG. 1 shows an overview of the entire transaxle of the presentinvention including an HST system. Referring also to FIGS. 2 and 3, thetransaxle is encased in an upper housing 1 and a lower housing 2 whichare secured by a plurality of bolts 145 and a liquid gasket seal 82 atthe joining surface of housings 1 and 2. Input shaft 75, which has alongitudinal axis 75A, extends through shaft opening 116 and issupported by bearing 7 and ring 5, which are retained by seal 4. Inputshaft 75 is driven by a belt (not shown) which is powered by a verticalshaft engine (not shown). As shown most clearly in FIG. 3, the rotationof input shaft 75 rotates the cylinder block 14 a of pump 14 at thespeed of input shaft 75. Pump 14 is of conventional construction,containing a series of piston receiving chambers 146, each of whichmovably mounts a pump piston 13 and piston spring 12 in a directionaxial to cylinder block 14 a of pump 14.

Pump pistons 13 are powered by piston springs 12 against thrust bearing9, which, as is well known in the art, is rotatably supported inswashplate 10 by a standard bearing and bearing guide structure,including bearing 112. Swashplate 10 is itself supported in upperhousing 1 by bearing cradle 8, as shown in FIG. 4.

Thrust bearing 9 acts as a ramp against which pump pistons 13 arepressed. The rotation of pump 14 causes pump pistons 13 to travel up ordown this ramp, thus creating an axial motion for pump pistons 13.Swashplate 10 may be moved to a variety of positions on bearing cradle 8to vary volume of oil pumped, which ultimately varies the speed of motor27, as described herein.

Movement of swashplate 10 is accomplished by the user's manipulation oftrunnion shaft 15, which in turn moves bearing guide 18. As is known inthe art, trunnion shaft 15 is supported by journal bearing 17, which isretained by seal 16. For example, if thrust bearing 9 is perpendicularto input shaft 75 and thus perpendicular to the axial plane of pumppistons 13, there will be no point along thrust bearing 9 where pumppistons 13 are forced axially, thus resulting in no axial motion forpump pistons 13 and no oil flow between pump 14 and motor 27. Thisposition is effectively a “neutral” position for the HST, in thatrotation of input shaft 75 will not ultimately result in movement of thevehicle.

The operator may move swashplate 10 by adjusting trunnion shaft 15,which varies bearing guide 18, in one direction to create a “forward”ramp at thrust bearing 9, so that axial motion of pump pistons 13 forcesthe oil flow in one direction. The operator may also reverse the flow bymoving thrust bearing 9 to the opposite, or reverse, position. Thedetails, of the resulting oil flow through the porting system of the HSTare set forth herein.

FIGS. 5 and 6 show center section 74 of the HST, which is securelymounted to upper housing 1 through bolt openings 103. Pump 14 isrotatably mounted on pump running surface 130 with center opening 138corresponding to shaft opening 116 to receive input shaft 75.

Motor 27 is rotatably mounted on motor running surface 61 byconventional means and is supported by motor shaft 22. When the HST isnot in operation, motor 27 is sealed to motor running surface 61 throughthe force of motor piston springs 25 against motor pistons 26, whichpress against thrust bearing 23 to create this seal. When the HST is inoperation, there is an additional force resulting from the oil pressure.Specifically, the interior of motor piston chamber 147 is sufficientlylarge enough that the flow of oil through passage 102 creates aresultant net balance of oil pressure in cylinder block 27 a in thedirection towards motor running surface 61, creating a seal at thispoint. Pump 14 is retained on pump running surface 130 in a similarmanner.

Center section 74 includes bearing structures 74A and 74B, which areintegrally formed therewith and include bearing openings 88 and 89.Motor shaft 22, which has a longitudinal axis 22A, is installed throughand fully supported by openings 88 and 89 and running surface 140. Themeans of supporting motor shaft 22 is a significant improvement over theprior art, which discloses the motor shaft supported at one end in thecenter section, and at the other end on some other external bearinghousing. The present invention eliminates the need for such anadditional bearing housing for motor shaft 22, reducing manufacturingexpense and weight, as well as reducing the overall size of the unit.

Proper alignment of motor shaft 22 is critical to the performance of theHST. The design of the present invention eliminates the necessity ofaligning such an additional bearing support with the support on centersection 74, resulting in an overall savings in weight and expense, aswell as increasing the ease of manufacture of the transaxle.

As shown in FIGS. 2 and 3, the longitudinal axis 22A of motor shaft 22is located between a first plane P1 formed at one end of input shaft 75and a second plane P2 formed at the opposite end of input shaft 75,where planes P1 and P2 are generally perpendicular to the longitudinalaxis 75A of input shaft 75. Similarly, the longitudinal axis 75A ofinput shaft 75 is located between a third plane P3 and a fourth planeP4, wherein planes P3 and P4 are generally perpendicular to thelongitudinal axis 22A of motor shaft 22, and plane P3 is formed at oneend of motor shaft 22 and plane P4 is formed parallel to plane P3 and atthe opposite end of motor shaft 22. It can also be seen in FIG. 2 thatthe two ends of motor shaft 22 are on opposite sides of a plane formedby the longitudinal axis 75A of input shaft 75 and perpendicular tolongitudinal axis 22A. Similarly, as shown in FIG. 3, the two ends ofinput shaft 75 are on opposite sides of a plane formed by longitudinalaxis 22A of motor shaft 22 and perpendicular to longitudinal axis 75A.

As most clearly shown in FIGS. 2 and 9, motor 27 also contains aplurality of piston chambers 147, each of which contains a motor piston26 and piston springs 25. Each motor piston chamber 147 has a passage102 to receive oil flow from arcuate ports 106 and 107 on motor runningsurface 61 of center section 74.

Each motor piston 26 is driven by the oil flow received through arcuateports 106 or 107 in a direction axial to motor 27 and against thegenerally circular motor thrust bearing 23. As shown in FIGS. 3 and 9,motor thrust bearing 23 is fixed in its position relative to motorpistons 26 at an angle such that the action of motor pistons 26 againstthrust bearing 23 creates a rotational movement of cylinder block 27 aof motor 27. Motor thrust bearing 23 is of standard construction and iscomposed of bearing plates 23 a and 23 b and bearing race 23 c. Motor 27is supported on and drives motor shaft 22. Cylinder block 27 a of motor27 has internal gear teeth (not shown) which mesh with gear teeth 45 onmotor shaft 22 to rotate motor shaft 22 at a speed equal to the rotationof cylinder block 27 a of motor 27.

A major improvement that this invention presents over the prior art isthe elimination of the need for an intermediate support for motor thrustbearing 23. As shown in FIGS. 1, 2 and 9, motor shaft 22 does not extendthrough the center of thrust bearing 23. Therefore, thrust bearing 23 isfully supported at its proper angle by upper housing 1 without the needfor an additional structural member such as is used for pump thrustbearing 9, which must be supported by swashplate 10. This results in aless expensive and simpler unit to manufacture, and the absence of theadditional member reduces the overall size and weight of the transaxleunit.

As described below, oil flow from pump 14 to motor 27 is the means bywhich rotational power is transmitted by the HST. Arcuate ports 136 and137 on pump running surface 130 provide the means for transferring oilfrom passage 101 of pump piston chamber 146 through oil passages 104 or105 and to motor 27. Arcuate ports 106 and 107, which are located onmotor running surface 61 and which coact with passages 102 of motorpiston chamber 147, act to receive the oil from oil passages 104 or 105and return it to pump 14.

It is to be understood that there are a plurality of pump pistons 13 andmotor pistons 26 and their related parts and chambers, and, therefore,the discussion herein of these parts in a singular sense is forconvenience only, and should not be read to limit the invention in anyway. In the preferred embodiment, there are five (5) pump pistons andseven (7) motor pistons.

As shown in FIG. 3, each pump piston chamber 146 has a passage 101opening for coaction with arcuate ports 136 and 137 on pump runningsurface 130 of center section 74. In the “forward” oil flow directiondescribed above, the oil flow created by the movement of pump pistons 13moves through passage 101 to arcuate port 137, and then through oilpassage 105 to arcuate port 106 on motor running surface 61, and finallyto passage 102 of motor piston chamber 147. The oil then returns to pumppiston chambers 146 through passage 102, arcuate port 107, oil passage104, arcuate port 136 and passage 101.

In the “reverse” oil flow direction described above, the oil essentiallytravels in a reverse direction, being forced by pump piston 13 throughpassage 101 and arcuate port 136 to oil passage 104 and arcuate port 107and passage 102, and finally to motor piston chamber 147. The oil isthen returned to pump 14 through arcuate port 106, oil passage 105 andarcuate port 137. The rotational direction of motor 27 depends uponwhether this oil flow is in the “forward” or “reverse” direction, asthis rotation, and ultimately the movement of the vehicle, will also be“forward” or “reverse.”

As can be seen in FIG. 2, the transaxle design includes expansionchamber 121 formed by external wall 3 and internal wall 124. Suchexpansion chambers are well-known in the prior art and are used toprovide a space for the oil to expand into during operation of thetransaxle. Expansion chamber 121 may be located at different areas alongthe upper and lower housings 1 and 2, and, in the preferred embodiment,expansion chamber 121 is located along upper housing 1 or lower housing2 outside differential gear 63.

As shown in FIGS. 1 and 2, braking for the transaxle is accomplishedthrough a braking mechanism 109 located on, and supported by motor shaft22 and comprising brake stator 57 and brake rotor 58, triggered by brakearm 53 and brake actuator 55. Braking mechanism 109 is located within acavity 110 which is separated from transmission cavity 48 by a standardseal 31.

The novel brake feature of this HST is clearly shown in FIG. 19.Specifically, FIG. 19 is a cutaway portion of the top view of a portionof the transaxle generally shown in FIG. 1, but including the novelbrake feature. The remaining elements of the transaxle shown in FIG. 19can be the same as those shown in FIG. 1.

Motor shaft 222, which can be identical to motor shaft 22 previouslydescribed, has, at one end, gear teeth 223 integrally formed thereon.Brake mechanism 250 includes brake rotors 258, which are rotatablymounted on gear teeth 223 of motor shaft 222 such that brake rotors 258rotate with motor shaft 222, and brake stators 257, which do not rotate.FIG. 19 shows lower housing 202 of the transaxle, which can otherwise beidentical to lower housing 2 previously described. Brake arm 253 isconnected to lower housing 202 through bolt 254, washer 255 and nut 256.When the brake mechanism 250 is to be activated, the user moves brakearm 253, which causes pins 259 to move in a lateral direction towardsbrake stator 257. This movement of pins 259 moves stators 257 intocontact with rotors 258, causing contact and friction between stators257 and rotors 258 and thus effectuating braking. Pins 259 are notmounted in a separate housing but are instead contained and held inmating half-round sections formed into both lower housing 202 and theupper housing (not shown) of transaxle. The advantage this designpresents is the elimination of separate housing elements for the pins,reducing the weight and cost of the unit.

As is known in the prior art, the present invention uses a differentialto transfer power from motor shaft 22 to the pair ofoppositely-extending axle shafts 62 and 62′ which are used to drive thevehicle. As shown in FIGS. 1–3, motor shaft 22 contains a center portion46 which contains gear teeth 126 which mesh with teeth 63 b ondifferential gear 63. Differential gear assemblies known in the artgenerally include an internal cross-shaft that serves as the actualdriving mechanism for the output axles. A key improvement in thisinvention is the use of a novel structure which eliminates the need forsuch an internal cross shaft on differential gear 63.

As shown in FIG. 1, the various differential gears are contained indifferential housing 64, which includes two identical opposing endcaps108 and 108′. Endcaps 108 and 108′ are shown in detail in FIGS. 10, 11and 12. Axle shaft opening 152 is integrally formed therein to receiveaxle shaft 62 or 62′. Bolt openings 154 and 154′ are also formed thereinto receive and secure bolts 68 and 68′.

As is shown most clearly in FIG. 3, planet gear 66 is mounted onto theinside of differential gear 63 through opening 63 c by means of a key orraised portion 66 a which fits into keyway 63 a formed in differentialgear 63. Planet gear 66′ is similarly located. Planet gears 66 and 66′are thus held in place by keyways 63 a and 63 a′ and endcaps 108 and108′. This arrangement replaces the cross-shaft of prior art designs,where the cross-shaft was used to support the planet gears.

Planet gears 66 and 66′ include a plurality of teeth 66 b and 66 b′,which are meshed with and cause the rotation of bevel gears 65 and 65′.Bevel gears 65 and 65′ are meshed with respective axle shaft gears 47and 47′ to cause rotation of axle shafts 62 and 62′.

Thus, each bevel gear 65 and 65′ is located and held in place by planetgears 66 and 66′ on one side and by endcap 108 or 108′ on the otherside. Endcaps 108 and 108′ function to center and hold bevel gears 65and 65′ and to allow the entire differential assembly to be heldtogether by two bolts and nut assemblies 68 and 68′. This is a much morecompact and less complicated design than has been used in the prior art.In addition, the elimination of a cross shaft removes the need for ahollow center section, thereby making the differential design of thepresent invention stronger than prior art models.

Another embodiment of this differential is shown in FIGS. 13–18, whereinplanet gears 266 are secured by and mounted on end caps 208.Specifically, as shown in FIG. 17, each planet gear 266 has a tab 267which may be integrally formed therewith, and tab 267 is mounted forrotation on curved mounting surface 215 on end cap 208. When two endcaps 208 and 208′ are mounted together as shown in FIG. 14, theirrespective mounting surfaces 215 combine to secure planet gears 266 inplace.

Each end cap 208 has a notch 220 which may be integrally formed thereinand which fits into keyway 268 formed into ring gear 263. As ring gear263 rotates, force is transmitted from the sides of keyway 268 to notch220 of end cap 208, causing the entire differential unit 200 to rotate.Thus each end cap 208 receives the rotational force of ring gear 263through notch 220 and transmits that force to planet gear 266, causingplanet gears 266 to move with the rotation of ring gear 263.

As shown in FIG. 14, the differential unit 200 is secured togetherthrough the use of a pair of bolts 275 and 275′ mounted through andsecuring end caps 208 and 208′. Each planet gear 266 engages and drivesbevel gears 265 to cause the rotation of bevel gears 265 about the sameaxis of rotation as ring gear 263. At the same time, each bevel gear 265engages and drives a rotatable output shaft 262 to power the vehicle inwhich the differential is used. Each bevel gear 265 has an opening (notshown) which corresponds to opening 271 on end cap 208, and which hasgear teeth (not shown) to engage and drive an output shaft 262, whichhas gear teeth 280 formed thereon. Bevel gears 265 are engaged on theinside of differential unit 200 by planet gears 266, and are engaged attheir outside surface 269 by mounting surface 270 on end cap 208. Eachend cap 200 has a shaft opening 271 which corresponds to bevel gearopening 302 to receive output shaft 302.

As discussed above, end caps 208 and 208′ may be bolted to one anotherusing bolts 275 and 275′ through bolt holes 301 to form a singledifferential unit. It is also possible to use one larger end cap inplace of the two separate end caps. In such embodiment the one large capunit is bolted to an outside face of ring gear 263 and holds androtatably mounts both planet gears 266.

The embodiment shown in FIGS. 13–18 shows the differential unit beingmounted within the center, i.e., between the outside faces of ring gear263. However, it is also possible for the planet gears 266, bevel gears265, and end caps 208 to be mounted off-center, such as on the outsideface of ring gear 263, with rotational force still being transferredfrom ring gear 263 to planet gears 266 through the single end cap unitsecured to ring gear 263 or through a set of end caps similar to thosedescribed above.

As shown in FIGS. 3, 7 and 8, center section 74 contains a check valvemechanism including check valve plate 41, ball 39 and spring 40. Plate41 is formed of powdered metal which is significantly harder than thecast aluminum used to form center section 74. Bottom face 79 of centersection 74 is shown in FIG. 6. Plate 41 is mounted on bottom face 79 bythree bolts 42 through bolt openings 127 and received by openings 128 onbottom face 79 of center section 74.

Plate 41 has top surface 148, which is flush with bottom face 79 ofcenter section 74 when mounted, and bottom surface 149. As shown in FIG.8, bottom plate surface 149 has generally circular opening 133 formedtherein, while top plate surface 148 has a slightly larger opening 131formed therein. Openings 133 and 131 coact with each other and withvalve openings 156 on bottom face 79 to form check valve 139. Checkvalve 139 includes ball support surface 135 to support ball 39 whencheck valve 139 is in the closed position, as shown in FIG. 3. When thecheck valve 139 is opened, ball 39 lifts off of ball support surface 135to allow oil from sump 155 to flow through check valve 139. Oil filters43 are used to prevent contaminants from entering sump 155 fromtransmission cavity 48.

A raised annular surface or ring 129 is formed around opening 133 on topsurface 148 of plate 41, and is pressed into the lower strength bottomface 79 of center section 74 to form a seal between plate 41 and centersection 74. The minimal leakage which may occur due to deflection in themetal does not affect operation of the transaxle because center section74 is within main transmission cavity 48, which is filled with oil.Thus, the present invention provides a simple, low cost sealingmechanism which allows for the use of a lighter cast aluminum centersection without the need for the use of additional machining to use anO-ring, as is done in the prior art.

Prior art HST designs have the pump and motor mounted either at a90-degree angle or in a parallel arrangement, whereby the pump and motorare set “back-to-back.” In the present invention, the preferredembodiment calls for these elements to be positioned on center section74 at the standard 90-degree angle to one another, as shown in thedrawings. However, if necessary, center section 74 could provide formotor running surface 61 to be inclined upwardly or downwardly in thevertical plane of FIG. 2. Such an orientation, which may be required bythe configuration of the vehicle, would also require motor shaft 22 toremain parallel to motor 27. In this position, motor shaft 22 is nolonger perpendicular to axle shafts 62 and 62′ and differential gear 63,as is required to have gear teeth 63 b and gear teeth 126 of motor shaft22 to mesh using standard gearing.

To allow such an arrangement, the present invention would require theuse of a helical gear at motor shaft center portion 46 or ondifferential gear 63 to allow these gears to properly mesh. Such helicalgears are well-known in the art, but have not previously been used inHST designs to allow the pump and motor to be oriented at angles otherthan the standard 90-degrees. The angle of the helix on such a gear isdetermined by the angle between the motor shaft 22 and the axle shafts62 and 62′.

With a transaxle, it is necessary to reduce the rotational speed of theinput shaft as it is transmitted to the final drive axles. One of thedisadvantages of prior art transaxle designs is the need to provide areduction of angular shaft speed through mechanical gearing. Suchmechanical reduction requires the use of extra gears, shafts, supportsand various other related parts, as shown in prior art patents. Thisresults in additional expense in manufacturing as well as additionalweight in the transaxle. Furthermore, mechanical gears are subject tofailure if stressed sufficiently or repeatedly.

In the present invention, at least a portion of this shaft speedreduction is accomplished through the hydraulics. In a preferredembodiment, this is accomplished by internally sizing motor 27 at alarger capacity than pump 14. As an example of the preferred embodiment,it has been discovered that if the capacity of motor 27 is 21 cubiccentimeters (cc), while the capacity of pump 14 is 10 cc, a significantreduction in the speed of motor shaft 22 is achieved. With such sizingit has been found that the angular speed of motor shaft 22 is generallyreduced to about one-half of the angular speed of input shaft 75.

In light duty applications where the prior art would require a doublemechanical reduction, the present invention can eliminate this secondarymechanical reduction altogether. In heavy-duty applications which wouldrequire two or three mechanical reduction units, the present inventionmay only require a single secondary mechanical reduction unit. In eitherevent, the present invention results in a significant savings in size,weight and expense over prior art designs. This also results in animprovement in reliability, as a hydraulic reduction is less susceptibleto breakdown due to the fewer number of moving parts required.Furthermore, a hydraulic reduction is less likely to break from beingoverstressed than is a mechanical gear reduction.

As seen in FIG. 1, axle shafts 62 and 62′ extend from differentialhousing 64 through transmission cavity 48 and axle cavity 49 to outeraxle bearings 72. The wheels (not shown) of the vehicle are thenattached at axle ends 150 and 150′. In prior art models, oil extendsthroughout axle cavity 49 along the length of the axle shafts tolubricate outer axle bearings 72 and is sealed in cavity 49 at seal 120.

However, inherent in the manufacture of any such axle shaft is a slightdeviation from the main axis at either end 150 or 150′ of axle shafts 62and 62′. Such minor deviations occur through imperfections in themanufacturing process and do not affect performance of the axle shaft orthe transaxle. Further deflection occurs due to axle loading at ends 150and 150′. The sum of these deflections together with any wear at theouter axle bearings 72 can create minor gaps at seal 120, which cancause leakage of oil from axle cavity 49. Such a gap at seal 120, andsubsequent oil leakage, can also occur through normal wear and tear.Wear of seal 120 and outer axle bearings 72 can cause detritus from theseal, bearing and surrounding structures to contaminate the oil.

In the prior art, oil leakage has been dealt with through the use ofextra bolts and sealant at the location of seal 120 as well as atadditional locations along sealing surface 125. This results inadditional parts, expense and weight for the unit.

Since the present invention does not fill axle cavity 49 with oil, thisproblem is eliminated without the need for such extra bolts or sealant.As shown in FIG. 1, seals 71 are used to prevent oil from flowing fromtransmission cavity 48 to axle cavity 49. Seal 71 thus operates as theprimary oil seal for transmission cavity 48 and, in a preferredembodiment, seal 71 is a seal made of nitryl.

In the present invention, a conventional higher viscosity grease withinaxle cavity 49 provides the necessary lubrication to outer axle bearings72. Use of this higher viscosity grease provides better lubrication tothe outer axle bearings 72 than is available through the use of oil.Seals 120 serve to maintain this higher viscosity grease within axlecavity 49 and thus do not serve as the primary oil seal. Moving theprimary oil seal from outer axle bearing 72 to seal 71 eliminates orminimizes oil leaks, extends the life of the product and reduces thequantity of oil needed in the casing. Seals 120 further act to minimizethe amount of outside contaminants which reach outer axle bearings 72.

Another important and novel feature of this invention is the hydraulicbypass shown in FIGS. 2 and 9. The effect of this bypass system is toenable the vehicle user to roll or “freewheel” the vehicle withoutresistance from the oil in the HST. When an HST does not have any powerbeing applied to it through the tractor motor, pump 14 and motor 27 arenot being rotated. Therefore, any attempt to roll the vehicle wouldtransmit the rotational energy through axle shafts 62 and 62′, andthrough differential gear 63 to motor shaft 22. This in turn will rotatemotor 27, and the action of motor pistons 26 against motor thrustbearing 23 causes axial motion of motor pistons 26, causing oil flowthrough the porting of center section 74. However, with pump 14 atneutral there is no place for the oil to go, and high pressure results.This high pressure causes resistance to further motion of motor shaft 22and axles 62 and 62′ and prevents the user from pushing the tractor.

Prior art solutions to this problem generally involve placing a valvebetween arcuate ports 106 and 107 to allow the oil to flow between thesetwo ports, i.e., between the pressure side and vacuum side of HST centersection 74. However, such a hydraulic valve allows only a limited amountof oil to pass between the ports due to inherent design limitations,such as the diameter of the hydraulic valve through which the oil isdiverted. Such a valve also requires accurate machining to maintainminimum clearances to reduce leakage during normal operation of theunit.

The present invention solves this problem by use of a mechanism to liftmotor 27 off of motor running surface 61 of center section 74, thusbreaking the seal at that point and allowing oil to flow out of arcuateoil port 106 and into transmission cavity 48. Thus, the oil is notported from the pressure side to the vacuum side, but rather bypassesthis entire circuit within center section 74.

To activate this feature, bypass arm 50 is manipulated by the user torotate bypass actuator 29. Seal 28 is used to retain oil within the maintransmission cavity 48 at this point. Bypass actuator 29 includes rod115, which is shaped at its base so that rotation of rod 115 forcesbypass plate 30 to press against the base of motor 27, breaking its sealto motor running surface 61. This allows the oil to flow from arcuateport 106 to transmission cavity 48. The oil is then returned to motor 27through arcuate port 107. This design enables the vehicle to readily“free wheel” with less resistance from the oil.

Further manipulation of bypass arm 50 and rod 115 causes bypass plate 30to withdraw off of motor 27, allowing motor 27 to return to its normalposition on motor running surface 61, reestablishing the seal at thatpoint. The design of the present invention could also be used in adifferent embodiment to lift pump 14 off of pump running surface 130, asthis would have the same effect.

An advantage of this design is that it is very simple and inexpensive tomanufacture and install because it does not require precise tolerances.Prior art hydraulic bypasses using valves to move the oil between itsporting sections require very precise machining of the valves to preventunwanted leakage, and are therefore more expensive to manufacture. Inaddition, this mechanism dissipates the oil into the cavity rapidly toallow immediate movement of the vehicle.

The above descriptions are intended to illustrate the various featuresof this invention and are not intended to limit it in any way. Furtheradvantages will be obvious to one of ordinary skill in the art. Thisinvention should be read as limited only by the following claims.

1. A center section for a hydraulic drive apparatus, comprising: a firstrunning surface comprising a first pair of arcuate ports formed thereon;a second running surface perpendicular to the first running surface andcomprising a second pair of arcuate ports formed thereon; a first systemport formed in the center section and connecting one of the first pairof arcuate ports to one of the second pair of arcuate ports; a secondsystem port formed in the center section and connecting the other of thefirst pair of arcuate ports to the other of the second pair of arcuateports; and a first check valve located in the center section andconnected to the first system port and a second check valve located inthe center section and connected to the second system port, each checkvalve comprising a valve body located in a check valve opening formed onan external surface of the center section; wherein at least a portion ofat least one system port and the entire body of at least one of thecheck valve bodies is located directly below the first running surface,whereby a line perpendicular to and through the first running surfaceintersects both the one check valve body and the portion of the systemport.
 2. A center section as set forth in claim 1, wherein at least aportion of the other system port and the entire body of the other checkvalve body is located directly below the first running surface, wherebya second line parallel to the first line and through the first runningsurface intersects both said portion of the other system port and theother check valve body.
 3. A center section as set forth in claim 2,wherein both of the check valves are oriented generally perpendicular tothe first running surface.
 4. A center section as set forth in claim 3,wherein the first running surface is generally horizontal.
 5. A centersection as set forth in claim 4, wherein the second running surface isgenerally vertical.
 6. A center section for a hydraulic drive apparatus,comprising: a first running surface comprising a first pair of arcuateports formed thereon; a second running surface perpendicular to thefirst running surface and comprising a second pair of arcuate portsformed thereon; a first system port formed in the center section andconnecting one of the first pair of arcuate ports to one of the secondpair of arcuate ports, wherein at least a portion of the first systemport is located directly below the first running surface; a secondsystem port formed in the center section and connecting the other of thefirst pair of arcuate ports to the other of the second pair of arcuateports; a first check valve connected to the first system port, the firstcheck valve comprising a first valve body located in a first check valveopening formed on an external surface of the center section, wherein atleast a portion of the first valve body is located directly below one ofthe first pair of arcuate ports; and a second check valve connected tothe second system port, the second check valve comprising a second valvebody located in a second check valve opening formed on the externalsurface of the center section.
 7. A center section as set forth in claim6, wherein at least a portion of the second system port is locateddirectly below the first running surface, and at least a portion of thesecond valve body is located directly below the other of the first pairof arcuate ports.
 8. A center section as set forth in claim 6, whereinboth of the check valves are oriented generally perpendicular to thefirst running surface.
 9. A center section as set forth in claim 8,wherein the first running surface is generally horizontal and the secondrunning surface is generally vertical.
 10. A center section as set forthin claim 9, wherein the external surface of the center section is formedopposite to the first running surface.
 11. A center section for ahydraulic drive apparatus, comprising: a pump running surface formounting a rotatable pump cylinder block; a motor running surface formedperpendicular to the pump running surface for mounting a rotatable motorcylinder block; hydraulic porting formed in the center section andcomprising first and second system ports, first and second arcuate portsformed on the pump running surface and third and fourth arcuate portsformed on the motor running surface, wherein the first and third arcuateports are open to the first system port and the second and fourtharcuate ports are open to the second system port; and a first checkvalve located in the center section and connected to the first systemport and a second check valve located in the center section andconnected to the second system port, each check valve comprising a checkvalve body located in a check valve opening formed on an externalsurface of the center section, wherein at least a portion of one of thesystem ports is located between the pump running surface and one of thecheck valve openings and the entire body of at least one of the checkvalve bodies is located directly below the pump running surface.
 12. Acenter section as set forth in claim 11, wherein both of the checkvalves are oriented generally perpendicular to the first runningsurface.
 13. A center section as set forth in claim 12, wherein the pumprunning surface is generally horizontal and the motor running surface isgenerally vertical.
 14. A center section as set forth in claim 13,wherein the external surface of the center section is formed opposite tothe pump running surface.
 15. A center section as set forth in claim 11,wherein both check valve bodies are located entirely below the pumprunning surface.
 16. A hydraulic drive apparatus, comprising: a housingforming a sump and having a center section mounted therein; a pump inputshaft extending into the housing and drivingly engaging a rotatable pumpcylinder block located on a horizontal pump running surface of thecenter section; a rotatable motor cylinder block located on a verticalmotor running surface of the center section and driving a motor outputshaft; the center section further comprising: first and second systemports located in the center section, wherein at least a portion of thesystem ports is located below the pump running surface; first and secondarcuate ports formed on the pump running surface and third and fourtharcuate ports formed on the motor running surface, wherein the first andthird arcuate ports are open to the first system port and the second andfourth arcuate ports are open to the second system port; and a firstcheck valve located in the center section and connected to the firstsystem port and a second check valve located in the center section andconnected to the second system port, each check valve comprising a checkvalve body located in a check valve opening formed on an externalsurface of the center section, wherein at least a portion of one of thesystem ports is located above one of the check valve openings and theentire body of at least one of the check valve bodies is locateddirectly below the pump cylinder block.
 17. A hydraulic drive apparatusas set forth in claim 16, wherein the entire bodies of both of the checkvalve bodies are located directly below the pump cylinder block.